For reasons of comfort and performance automated transmissions that can be powershifted are increasingly used in utility vehicles. In this context double clutch transmissions that shift without traction force interruption are particularly attractive. Such automated change-speed transmissions comprise an intermediate transmission or a number of intermediate transmission groups connected one after another, and if necessary a downstream planetary transmission. The gears in a classical double clutch transmission are divided into two transmission branches, a direct and an indirect gear group, wherein each group is associated with one clutch of the double clutch. With the help of the double clutch a sequential shift sequence almost free from traction force interruption can be carried out, wherein in each case the next gear is preselected in the currently load-free transmission branch and the gearshift takes place by overlapped actuation of the two clutches. The force flow of the gears can pass conventionally by way of a drive input shaft and a drive output shaft, or it can meander through the transmission via a plurality of changing shafts.
Compared with a pure powershifting automatic transmission of planetary design, automated transmissions have efficiency and cost advantages. However, as the number of gears increases so too do the structural size, the design complexity and hence the manufacturing costs. Since, depending on their field of use, utility vehicles as a rule need a relatively large number of gears in order to achieve a given transmission spread and for efficient operation, it is expedient particularly for utility vehicle applications also to consider less expensive and more compactly configured mixed transmission forms, so-termed partial double clutch transmissions, in which besides a powershifting transmission or transmission section with a double clutch, a conventional transmission section, i.e. one which shifts with traction force interruption, for example a main transmission group or a downstream transmission that shifts with traction force interruption, such as a transfer box or axle transmission, is provided. Depending on the shift carried out, these partial double clutch transmissions behave like a conventional change-speed transmission, i.e. with the disadvantage of a traction force interruption if the gearshift takes place in the main transmission section or like a powershift transmission if the gearshift takes place in the double clutch transmission section.
For example, DE 10 2008 008 496 A1 describes a multi-stage transmission which corresponds in function to a partial double clutch transmission. The multi-stage transmission has a first and a second input shaft which, by means of the respective clutches of a double clutch, can be connected to a drive engine, and a countershaft to which the two input shafts are coupled by way of a first or a second input gearset with different gear ratios. A drive output shaft can be coupled selectively via various other gearsets to the countershaft. A two-directional gearshift between a first and a second gear can be carried out without traction force interruption and without further shifting processes by overlapped opening and closing of the two clutches, since the force flow in those gears only changes between the first and second gearset but passes on to the drive output by way of the same other gearset. The other gearshifts require a shift of the gear clutches involved, with the consequence of a traction force interruption.
The comfort and performance deteriorations caused by a gearshift with traction force interruption are made even worse by the loss of supercharge pressure in turbocharged internal combustion engines. Exhaust gas turbochargers are used with most diesel engines and more recently in many Otto engines as well. An exhaust gas turbocharger has a turbine which makes use of the kinetic energy contained in the exhaust gas for driving a compressor which, for its part, draws in fresh air and passes it, after pre-compression, to the cylinders of the engine to boost its power. It is therefore driven by the exhaust gas flow, without any direct relationship to the engine speed. With a correspondingly high driving resistance, in addition to the speed loss caused by shifts with traction force interruption, insufficient turbocharging brings about a loss of traction power which, when a powerful acceleration is required in the lower engine speed range, is generally referred to as turbo lag. Thus, if the supercharge pressure of the supercharger first has to be built up, the suction torque needed for reaching the full-load torque of the engine is only available above a limiting supercharging speed. Since the quantity of air delivered is approximately proportional to the square of the supercharging speed, the time interval after an upshift under load between reaching the suction torque and the engine torque attainable when stationary is clearly perceptible as a traction power loss. Particularly in the case of utility vehicle diesel engines this time interval, and thus the transition into the supercharged operating range, is often particularly long.
When a shift process takes place in the conventional section of the transmission, partial double clutch transmissions are as much affected by this problem as any conventional automated transmission. To avoid turbo lag, it is true that electrical or mechanical auxiliary drives of the exhaust gas turbocharger or auxiliary compressors are known, which are designed to increase the supercharge pressure if the engine is providing insufficient supercharging energy. These devices, however, are relatively expensive and take up additional structural fitting space.
From the previously unpublished patent application DE 10 2010 028 076.3 a method is known for controlling shifts in an automated change-speed transmission with an upstream hydrodynamic starting and braking element, by means of which during a traction shift the occurrence of a traction power loss of a turbocharged drive engine while the load is building up after the shift, can be avoided. In this case a hydrodynamic torque converter with a pump impeller wheel, a turbine wheel and a guide wheel, or a hydrodynamic clutch in the absence of a guide wheel, is provided, wherein by way of a bridging clutch the pump impeller can be connected to an intermediate shaft itself connected to a transmission input shaft, and the turbine wheel can be connected to the intermediate shaft by a freewheel clutch and can be braked relative to a fixed housing by means of a turbine brake. The load is reduced for disengaging the gear at the beginning of the traction shift while maintaining the engine torque delivered by the drive engine by building up a correspondingly high resistance torque, in that the turbine brake is at least partially closed and the torque transmitted by the starting and braking element to the turbine brake is adjusted to the engine torque of the drive engine. In this way the gear can be disengaged while largely free from torque. After the synchronization and engagement of the target gear, the turbine brake is opened again so that the load corresponding to the new gear is applied to the engine. Since the engine torque is not reduced in order to decrease the load, the disturbing drop of the engine torque delivered spontaneously under load by the turbocharged drive engine does not take place.
In the known method the effect of turbo lag is reduced by means of a hydrodynamic converter or its turbine brake. But in a partial double clutch transmission no such starting and braking element is present, but instead, a double clutch is provided. Accordingly the known method cannot be applied to the problem of turbo lag during a shifting process in a partial double clutch transmission.